The useful work of internal combustion engines is limited by their mechanical efficiency. On average, about 85% of the work available on the piston at full load is available as useful work on the flywheel due to internal engine friction. At lower loads, the above figure is even lower. Piston assembly friction (piston and piston rings) alone can account for up to 75% of the overall mechanical losses. Thus, piston friction reduction is highly desirable. Furthermore, engine parts are subject to wear that eventually limits the engine power and efficiency as well as increasing oil consumption, and increasing exhaust emissions. The most critical wear occurs on the piston rings and cylinder liners. Excessive wear requires engine overhaul or replacement. Thus, liner and ring wear reduction is also highly desirable.
Piston rings have two primary functions: Limiting oil flow into the combustion chamber and minimizing blowby (leak of high pressure combustion gas from the combustion chamber into the crankcase). Both functions are accomplished as high pressure combustion gases force the rings against the cylinders and the lower part of the piston groove and thereby seal the relatively large clearance between piston and liner. The top ring is subject to the highest pressure loading and thus suffers the most wear and has the largest contribution in friction.
The "rotating sleeve engine" is an invention that can significantly improve the lubrication conditions of the piston and piston rings, eliminate or significantly reduce wear and significantly reduce piston assembly friction.
In reciprocating piston engines, the piston linear speed is reduced to very low values at the regions with proximity to top and bottom dead centers. In those parts of the stroke, the sliding speed between the compression rings and the liner is insufficient for the maintenance of hydrodynamic lubrication. The protective lubricant film gradually breaks down and metal to metal contact occurs. The high cylinder pressure during the compression and power strokes loads the compression rings further, intensifying the phenomenon and expanding the portion of the stroke where the metal to metal contact occurs. Thus, localized wear on the liner around the dead centers and especially at the top is typical after prolonged engine operation. At the regions around the mid portion of the stroke, the piston speed reaches sufficient values for the hydrodynamic lubrication regime. The protective lubricant film prevents metal to metal contact, reduces the friction coefficient by up to two orders of magnitude, and essentially eliminates wear. This can be verified by the fact that the mid portion of the liner is always free of wear. Numerous frictional experiments reveal increased piston assembly friction around the dead centers due to the described phenomenon.
The above phenomenon is further illustrated by the Stribeck diagram shown in FIG. 1 as presented by Irving J. Levinson, Machine Design. This diagram shows the friction coefficient between two sliding surfaces in the presence of lubricant as a function of the "duty parameter" which is defined as the product of sliding speed and lubricant viscosity divided by the normal contact pressure of the surfaces. When two surfaces slide in the presence of lubricant, three possible modes of lubrication are possible. At very low sliding speed and high normal load, boundary lubrication is present. Metal to metal contact is unavoidable. Due to surface adhesion, high level of friction and wear is present. As sliding speed and thus the duty parameter increases, hydrodynamic oil film pressure builds up, supporting a larger portion of the normal load. Thus, the two surfaces are gradually separated by the oil film with less and less asperity contact and reduced adhesive wear (mixed regime). Finally, at higher sliding speeds (duty parameter values of 50 or higher according to the Graph 1), the hydrodynamic pressure supports the entire load resulting in full separation. The metal to metal contact as well as wear are eliminated. In the part of the cycle when the piston approaches a dead center, the sliding speed approaches zero. Furthermore, when the piston is in proximity to the top dead center, compression-expansion stroke, the high cylinder gas pressure increases the normal load between the liner and the piston rings (which are practically pressure activated sealing devices) further reducing the value of the duty parameter. The result is that for a significant portion of the cycle, the duty parameter falls bellow the value of 50, with the corresponding high friction coefficient and level of wear. ##STR1##
Graph 1. Stribeck diagram
The cylinder liner (also called the "sleeve") of the rotating sleeve engine rotates with the objective of maintaining a non zero sliding speed and large values of duty parameter throughout the stroke. According to the Stribeck diagram, the friction coefficient is reduced by almost two orders of magnitude for that particular portion of the stroke. The rotation can be achieved via gear mechanisms from the crankshaft (similarly to a distributor or injection pump). For best results, the magnitude of the rotation needs to be high enough in order maintain the hydrodynamic lubrication regime between the compression rings and liner, even when the piston linear speed is zero and the cylinder pressure is at its maximum value.
In conventional engines, the rings must be free to rotate to minimize localized ring wear. However, both blowby and, for spark ignition engines, hydrocarbon emissions are affected by the relative azimuthal positions of the end gaps of the compression rings (Roberts and Matthews, 1996). When the rings are free to rotate, the engine designer cannot take advantage of these dependencies to help control blowby and hydrocarbon emissions. For the rotating sleeve engine, the rings can be pinned to prevent their rotation (which is no longer required to minimize wear).
In order to further investigate the feasibility of hydrodynamic lubrication just due to sleeve rotation, the Reynold's partial differential equation as shown by Hamrock (1994) was solved numerically in a situation that simulates a stationary piston ring subject to cylinder gas pressure while the liner rotates. The objective of the simulations is to explore the magnitude of the average hydrodynamic pressure obtainable by different liner sliding speeds and different ring profiles with a constant film thickness. The value of that pressure represents the maximum gas pressure that can be supported by the ring and still maintain the assumed film thickness. This pressure is the cylinder pressure at top dead center (TDC) compression stroke and is nearly equal to the peak cylinder pressure. The constant film thickness eliminates the contribution of squeeze film lubrication in the hydrodynamic film pressure and thus represents the worst case scenario for the rotating sleeve engine. It is as if the piston stays at top dead center indefinitely while the top compression ring is constantly loaded with high gas pressure The value for the lubricant viscosity was for a 20W oil as given by Hamrock (1984). This is a low viscosity lubricant that minimizes the hydrodynamic losses at mid stroke. A flat piston ring profile was assumed with surface irregularities as the only means for pressure build-up. This phenomenon is called "microhydrodynamic lubrication" Hamrock (1984). The surface irregularities were set equal to the combined surface roughness used by the ring-pack modeling performed by Tian and coworkers (1996) of 0.3 microns. The irregularities were assumed to be on the liner surface only (while the ring surface was assumed to be perfectly flat) and their shape was a 2 dimensional sinusoidal wave. With a 3 m/s liner sliding speed and a mean film thickness of 1 and 0.8 microns (within the range of hydrodynamic lubrication for this size of the combined asperity size according to Tian and coworhers) the average lubricant pressure was 16.03 and 28.27 atm respectively. Furthermore, the average pressure demonstrated an almost proportional variation with liner speed.
In order for the lubricant film of 0.8 microns minimum film thickness and with low viscosity oil to be able to support the typical peak pressure for spark ignition engines of about 50 atm, liner speeds of over 6 m/s are necessary. For heavy duty engines where the peak cylinder pressure can reach 100 atm, even higher speeds would be required. Note that if the liner speed is not sufficient for the peak cylinder pressure, the film thickness will further drop with some metal to metal contact at dead center. However, the boundary lubrication will be still confined to smaller part of the cycle, where the piston speed is nearly zero, and thus the energy losses due to boundary lubrication are still minimized.
In a preferred embodiment, a new ring profile is incorporated in order to enhance hydrodynamic pressure with lower liner speeds. In a conventional engine, the top compression ring is equipped with a barrel shape with the intention of creating converging surfaces (FIG. 6A) that enhance the build up of hydrodynamic lubricant pressure due to the up and down motion. Note that the depth of the barrel is ony 20 or 30 microns, and thus the profile looks perfectly flat to the naked eye. For the rotating sleeve engine, the depth of this shape can vary in the peripheral direction periodically as shown in FIG. 6B in order to create converging surfaces in the direction of sleeve rotation. Note that the shape is actually curved rather than angular as shown in FIG. 6. At a point of maximum depth, the barrel shape remains like a typical compression ring. However, this depth drops linearly with peripheral displacement, until the barrel shape is reduced to a perfectly flat surface. Then, suddenly, the barrel shape is reintroduced and the same process repeats periodically. The result is that multiple "wedges" or converging surfaces are formed that enhance hydrodynamic pressure due to sleeve rotation. The length of these wedges along the periphery of the ring range between 1 and 1.5 ring widths.
The Reynold's equation was solved again for the new profile. With a film thickness of 0.8 microns and a sliding speed of 3 and 4 m/s, the average lubricant pressure was 88.10 and 110.82 atm respectively. Smaller film thickness and higher sliding speeds yield even higher pressure. As discussed above, if the slight piston motion present around dead center and squeeze film lubrication are included in the problem, similar hydrodynamic film pressure and/or higher minimum film thickness can be achieved for the short period of time that piston sliding speeds are low and cylinder gas pressure is high, with lower liner rotational speeds.
The above simulations indicate that hydrodynamic lubricant film pressure can be created at TDC or BDC just by the surface irregularities, even for a moderate liner rotation (3m/s corresponds to 409 rpm for a 5.5 inch bore engine) and with a low viscosity lubricant. With the proposed ring profile, the lubricant pressure can be dramatically increased even with a relatively large film thickness reaching the magnitude of typical peak cylinder pressure for heavy duty diesel engines at full load. Note that the proposed ring profile increases the flat portion of the ring improving sealing and increasing the effectiveness of squeeze film lubrication at dead center. Part of the converging surface that enhances hydrodynamic lubrication due to up and down motion has been sacrificed. However, results from the models by Tian and coworkers (1996), Lawrence (1988) as well as several experimental studies indicate that there is more than sufficient film thickness at mid-stroke for hydrodynamic lubrication. Furthermore, since at least some liner rotation will be retained at mid-stroke, the converging surfaces at the peripheral direction will remain active and substitute for the overall reduction of the barrel shape.
Moving sleeves have been proposed for use in prior internal combustion engine patents. However, in the prior art, the objective of sleeve motion is replacement of the conventional poppet valves with intake and exhaust slots that are exposed by the motion of the liner. For example Giorgio in U.S. Pat. No. 5,482,011, discloses an engine design where a liner rotates inside the engine block. The liner is tightly fitted on the block and is provided with a port which is aligned with similar ports on the block for the intake and exhaust processes. The rotational speed of the liner is restricted to one half the crank speed due to timing requirements. Akira in U.S. Pat. No. 5,191,863 describes an engine design with a ported rotating liner used for intake and exhaust processes. Again, the liner's motion is restricted by timing.
Some of the existing moving sleeve engines can show some improvements in the piston lubrication in respect to a stationary liner when sleeve rotation occurs at the dead centers, even though that was not the objective of the invention. An example is the series of sleeve valve engines developed by Ricardo before WWII (U. S. patent unknown) which shows great similarity to the engines described by Giorgio and Akira. However, in all these designs, the sleeve motion can not be optimized for friction optimization due to timing restrictions. Furthermore, even though Ricardo and coworkers (1968) reported some potential benefits in that particular design in terms of piston friction, the tight tolerances between the liner and block throughout the liner external area (necessary mainly due to the sealing requirements of the port openings provided on both the sleeves and stationary cylinders) introduced large loads on the driving mechanisms, minimizing the potential frictional benefits.
The ported engine designs shown by Richardo and coworkers as well as other US patents are using oil for the lubrication of the outside sleeve surface. Relatively large quantities of oil can be expected to enter the intake and exhaust flows through the ports. However, the oil consumption and resulting hydrocarbon emissions will not be compatible with modem US regulations. If different means of lubrication of the sleeves of the ported engines is attempted in order to avoid the excessive oil consumption (i.e. solid lubricant or dry lubrication), the advantages on wear and low friction coefficient inherent in fluid lubrication will not be present. The present invention incorporates conventional valvetrain, which prevents oil from entering the flow of intake or exhaust gasses.
In order to further support the usefulness of the invention, a study in the scientific literature concerning engine lubrication has been performed and demonstrated in the following pages. Issues of the design and operation of sleeve valve engines as related to engine lubrication are reported as published in relevant literature. The sources of piston ring wear are further analyzed. Also, the frictional savings due to the elimination of the friction component due to metal to metal contact of the piston rings with the liner have been estimated. The additional friction due to rotation has also been estimated and when compared to the frictional savings, it is smaller.
Background on Moving Cylinder Sleeves
Moving sleeves are not a new or untried engine feature. Ricardo and Hempson (1968) describes in detail the highly successful "sleeve valve" engines developed in the period between WWI and WWII and during WWII, mostly for aircraft applications (spark ignited engines). The objective of that design was the replacement of the conventional valve train and poppet valves. The motion of the sleeve would expose intake and exhaust slots at the right time in the cycle. The main advantage for aero engines was the reduction of the frontal area of the engine by the elimination of the rocker and/or overhead cam mechanisms. The shape and motion of the sleeves was designed in order to optimize the port exposure. A crank rotating at half the engine speed was connected by a ball joint to the sleeve, causing it to reciprocate and twist (at top dead center compression stroke, the sleeve's motion was momentarily purely rotational). Other advantages of these engines included central spark plug placement for denotation reduction and volumetric efficiency improvements. A second set of rings was necessary in order to seal the sleeve-cylinder head gap. This ring-pack was installed on the cylinder heads and was stationary.
During the early 20's two experimental single cylinder engines were constructed in Ricardo's laboratory in order to further explore the potential of the sleeve valve concept. One was a conventional poppet valve engine with 4 valves, the other a sleeve valve engine. The two units were similar in every other respect. During the test, it was discovered that a "puzzling" feature of the sleeve valve engine was its lower frictional losses, in spite of the larger number of moving parts and large rubbing surfaces. Ricardo attributed this effect to the possibility of elimination of piston ring boundary lubrication due to the continuous ring-liner motion and continuous fluid lubrication of the rings. That theory was confined by the observation that the sharply localized wear, always found in the liners of poppet-valve engines at the point where the top piston ring comes to rest at top dead center, was absent in the sleeve valve. Later tests on large number of sleeve- and poppet-valve engines of various shapes and sizes indicated that the overall mechanical losses of the sleeve-valve engine were usually less than those of the poppet-valve. According to Setright (1975), the Bristol "Centaurus", a radial 18 cylinder air cooled high performance aero engine is holding the record for the longest operation between overhauls for a piston aero engine (3000 hours). Furthermore, the Nappier Sabre, a 24 cylinder liquid cooled high performance sleeve valve aero engine could maintain its combat rating almost indefinitely due to improved piston lubrication, while engines of the period with similar or lower Brake Mean Effective Pressure (BMEP) could maintain combat rating for only 5 minutes.
Evidence from the sleeve valve engines also indicates that the maintenance of hydrodynamic lubrication of the compression piston rings by the liner motion is feasible even with the relatively low rate of liner rotation that those engines had when the piston was at top dead center (TDC) compression-expansion stroke. Thus, wear protection of piston rings and liner is also feasible. Also, in the previous design, the primary function of the sleeve motion on the sleeve valve engines was not friction reduction. The clearance between sleeve and block was held low for gas sealing purposes around the intake and exhaust ports that were drilled on the sleeves and block. In the current invention, the sleeve- block clearance can be chosen for friction optimization. Drawings of the sleeve valve engines reveal large rubbing surface between the block and sleeve. The current invention reduces this rubbing area (by the introduction of the journal bearings) to the minimum necessary for fluid lubrication, and optimizes the sleeve motion for piston friction as well as sleeve friction. Thus, much higher frictional benefits are possible. Also, the reduction or elimination of metal to metal contact at TDC in those sleeve valve engines was possible due to the relatively low peak cylinder pressure typical of a spark ignition engine (compression ratio of about 6.5:1). In order to achieve the same effect in a turbocharged heavy duty diesel engine with a compression ratio of 15:1 or higher and peak cylinder pressure of about 100 atmospheres, higher rates of rotation and/or the proposed ring profile will be necessary.
Squeeze Film Lubrication
Lawrence suggests the possibility that "squeeze film" lubrication can protect the compression rings even at the critical moments when the piston approaches a dead center under certain operating conditions. Squeeze film lubrication is a tribological situation where, even though the sliding surfaces suddenly come to a complete stop, due to lubricant viscosity and inertia, the surfaces could still remain separated for a certain amount of time.
However, experimentation by Gauthier and coworkers (1987) on a diesel engine indicate that squeeze film lubrication can completely prevent metal to metal contact only if very high viscosity lubricant is used. However, when the lubricant viscosity is increased, hydrodynamic piston friction at mid stroke will increase dramatically, with a strong penalty in overall engine efficiency. Furthermore, typical lubricants cannot maintain that level of viscosity at operating temperature. Experiments by Mitsumoto and coworkers (1989) on diesel engines also show that increasing the lubricant viscosity improves the durability of the engine, with a penalty in efficiency.
Similar studies performed on spark ignited engines (Takiguchi et al., (1988); Ku et al., (1988); Ohmori et al., (1993)) show that squeeze film lubrication cannot prevent metal to metal contact of the compression rings under any operating condition. However, slight reduction of wear followed by an increase in overall friction was observed when the lubricant viscosity is increased.
The fact that squeeze film lubrication is possible under extreme conditions on diesel engines with higher viscocities, and unlikely in spark ignition engines is mainly due to the ring design differences. The design of diesel engines is dominated by durability requirements, and thus they are generally equipped with wider compression rings. The larger surface area will enhance the squeeze film lubrication. However, a wider ring will also exhibit higher hydrodynamic friction at mid stroke. With a rotating liner, dependence on squeeze film lubrication will be minimized. Therefore, the piston ring width can be reduced without any durability trade offs. Thus, the mid-stroke hydrodynamic friction of diesels (which seems to be relatively large in respect to spark ignition, as shown bellow) can be also reduced with the new concept.
Boundary Contribution
In the current invention, the friction reduction is achieved mainly by the elimination or severe reduction of the boundary contribution in piston ring friction. A comprehensive literature review was conducted in order to estimate the contribution of boundary lubrication in the total piston friction in different kinds of engines and operating conditions.
In order to illustrate the typical piston friction behavior and the boundary lubrication at the dead centers, the crank-angle resolved data by Ku et al. for a 4.1 liter Cadillac spark ignition engine at 2000 rpm, light load are presented in Graph 2. A complete thermodynamic cycle is shown (two complete crankshaft revolutions). The x-axis shows crank angle degrees (dead center at 0, 180, 360, 560, and 720). The y-axis show piston friction force (N) and piston speed (m/s*10). The spikes at top dead centers clearly indicate the existence of asperity contact and boundary lubrication. Please note that in many similar experiments on different engines and operating conditions, those spikes are more clearly defined. The doted line shows the calculated instantaneous piston speed. The flat line (that alternates values between 0 and 10) shows where the friction can be considered predominantly hydrodynamic (value of 0) and boundary (value of 10). ##STR2##
Graph 2. Piston assembly friction on a 4.1 liter Cadillac V8 engine
For different operating conditions, the boundary and hydrodynamic contributions are continuously varying. Clearly, the hydrodynamic contribution is increasing with speed. Patton et al. suggests a nearly linear variation of the hydrodynamic friction torque portion with engine speed for spark ignited engines. This is supported by a number of experimental results (Patton) as well as the Stribeck diagram. Patton also suggests that the hydrodynamic portion is not sensitive to engine load. The piston friction which is the largest contributor of the hydrodynamic portion due to its larger than the rings rubbing area could be affected by load by the reduction of the oil film thickness in the thrust side during the power stroke. However, the film thickness will increase on the anti thrust side, making the overall effect not as severe. From the Stribeck diagram, it can be seen that if in a situation well into the hydrodynamic regime the load increases, the friction coefficient drops. It is unclear what will happen to the friction force (which is the product of normal load and friction coefficient) but it is clear that the sensitivity is not as high as the boundary friction where the friction coefficient is flat (if well into the boundary regime) or rapidly increasing with load (if in the mixed regime where the friction coefficient could be very sensitive to the duty parameter, and thus load). Thus, it seems justified for Patton et al. to attribute the piston assembly fmep (friction mean effective pressure) increase due to increasing load entirely on increase of boundary friction. This may not be entirely accurate for reasons described above, but it seems like a reasonable assumption.
In contrast with spark ignition engines, the piston assembly finep for diesel engines does not always increase with increasing load. Ball et al. conducted friction experiments on two 1.6 liter automotive engines, one diesel and one spark ignition. Although in the second case, the friction increased with load, it decreased in the second. This was attributed to the fact that the lubricant viscosity on the cylinder walls dropped due to the increased temperatures at higher loads. Thus, even though the boundary friction increased due to the increased gas loads on the rings, the hydrodynamic portion dropped even more. Gauthier et al. also supports the competing effects of increasing temperature and gas loading. In their measurements, the total piston friction was insensitive to load. This effect was not as apparent in Marek et al.'s experimental results. In their case, the gas loading term prevailed and the friction increased with load (in spite of the fact that they held the oil sunp temperatures lower than normal operating conditions, and thus amplifying the hydrodynamic effect). At 980 rpm, the friction at full load was 31.4% higher than the motoring conditions. The different behavior of these two engines (Gauthier's and Marek's) is probably due to design differences of the two engines. For example Gauthier's engine had a larger stroke (110 mm compared to 95 mm) which resulted in higher mean piston speeds.
The reason that diesel engines are generally less sensitive than spark ignited is partly due to the fact that Diesel engines run unthrottled. Thus, during the compression stroke, the cylinder pressure versus crank angle trace does not change with load. On a spark ignition engine, the density of the charge drops at lower loads, and so does the compression pressure.
However, the design differences of diesel and spark ignition engines also contribute to the different frictional behavior. Generally, diesel engines have larger piston skirts and wider piston rings. The higher durability requirement of the diesel engines and the lower speed range seem to be the main reason for this difference. Furthermore, the stroke (as well as the stroke to bore ratio) is higher on diesels, which creates a higher mean piston speed. As a result of the above, the hydrodynamic contribution should be higher in diesels (for a given speed). Thus, a reduction of lubricant viscosity at increasing loads can have a significant effect on total frictional losses. In a spark ignited engine the changes in the hydrodynamic portion seem not to be sufficient to significantly impact the total losses and counteract the increase of boundary friction.
By varying certain parameters that directly affect the hydrodynamic portion, Gauthier at al. was able to calculate the boundary contribution at 1250 rpm for different lubricant viscosity for their engine. The approximate boundary contribution on piston friction and total piston friction mean effective pressure (FMEP) for 1250 rpm motored is shown in Graph 3 and 4 respectively. The small increase in FMEP at very low viscocities is due to the very rapid increase of boundary friction. In the viscosity ranges at operating temperatures (less than 10 mm.sup.2/ s) the boundary contribution ranges between 12 and 25%. At the viscosity of minimum friction (4 to 5 mm.sup.2 /s) the boundary contribution is over 15%. It can be expected that if this engine was firing and under significant load, the boundary contribution would have been higher. If a rotating liner is applied on this engine with the optimum lubricant viscosity (about 5 mm.sup.2 /s), it is possible to eliminate that 20% boundary contribution (resulting from metal to metal contact between the compression rings at the dead centers). The rotating liner, apart from the continuation of the fluid lubrication of piston rings at dead centers, will also create extra hydrodynamic pressure on the piston skirt when the large thrust forces are applied by the connecting rod (similar to a journal bearing). Thus, the piston skirt surface area can be reduced without reducing the minimum film thickness between liner and piston and without increasing the chances of metal to metal contact between piston and liner (according to the Stribeck diagram, an increase in the load per unit area can be offset by the increase of the sliding speed keeping the value of the duty parameter unchanged). Thus, the reduction of the hydrodynamic friction of the reciprocating piston motion due to that skirt size reduction will overcome the increase of total friction caused be the relative rotary motion introduced between the piston and liner. Furthermore, the fact that the boundary term will be considerably reduced or eliminated, the FMEP will continue dropping with decreasing oil viscosity beyond the value of 5 mm.sup.2 /s. ##STR3##
Graph 3. Boundary contribution ##STR4##
Graph 4. Total friction
The friction of the engine used in Marek's experiment was far more sensitive to load than the one that Gauthier et al. used. Therefore, the boundary contribution in that engine is higher, with higher frictional benefits for the present invention. However, no effort was done to quantify the boundary contribution in that study. Needleman and coworkers expect large boundary contributions in diesel engines as well and suggest that "due to boundary lubrication, 40 to 50% of frictional losses of an engine are attributed to piston/ring assembly with 2/3 of the losses assigned to the top compression ring". In general, the frictional savings with the rotating sleeve concept applied on diesel engines could be higher than in the case of Gauthier's study.
As discussed earlier, due to different design trends and lower durability requirements, the boundary contribution in spark ignition engines can be expected to be higher. Patton et al. have developed empirical equations that estimate the piston assembly boundary and hydrodynamic friction contributions as a function of speed and load. By using the equations proposed by Patton, at 2000 rpm, the boundary contribution on total piston friction on a typical automobile engine can be calculated as high as 50% and approximately 20% at a speed of 6000 rpm at medium load and intermediate values in between.
Spark ignition automotive engines are required to operate over a wide speed range and the ringpack design is a compromise between wear and high speed friction. If the rotating sleeves are driven with some sort of gear mechanism, the optimization of the gear ratio is also a compromise. If the gear ratio between the crankshaft and the sleeves is selected so that the sleeve rotates at a sufficient magnitude for compete elimination of boundary friction at low engine speeds, excessive hydrodynamic friction could result at higher speeds. On the other extreme, if the gear ratio is such that just sufficient sleeve speed is present at high engine speeds, the sleeve speed at low crankshaft speeds will drop proportionally and may not be sufficient for complete elimination of the boundary/mixed friction at top dead center and high load when low viscosity oil is used. However, even in that case (some metal to metal contact occurs at top dead center compression stroke at high load and low engine speed) the portion of the stroke that this happens is confined to a much smaller part of the stroke around the dead center in respect to a conventional engine because the sliding speed of the compression ring is always held at well above the zero value. Since at these parts of the stroke the piston speed is very low, the piston energy losses due to boundary friction will still be far less than the in a conventional engine. Furthermore, due to the very large contribution of boundary/mixed friction at low speeds in conventional spark ignition engines, the frictional benefits could be very significant (especially at higher loads). At higher speeds when the piston and sleeve are moving faster, complete boundary elimination seems more likely but with a smaller potential for friction reduction.
In spite of the possible presence of some metal to metal contact at low speeds, some useful wear reduction can be possible in engines with a large speed range variation. Ohmori and coworkers (1993) showed that at 6000 rpm the instantaneous ring wear could be an order of magnitude higher than at 2000 rpm. Thus, even if some boundary friction still exists at low speed, the resultant wear will not be so significant. The more significant wear rate normally present at higher engine speeds can be eliminated.
Friction Calculations
The following calculations apply to a spark ignition automotive engine. The friction model developed by Patton et al. was used. The driving mechanism considered was the one of alternating sleeve speed (high magnitude when piston close to a dead center, low when at mid stroke, no reversal of direction considered, see embodiments). Note that in a 4 cylinder engine where all pistons reach a dead center at the same time, only one alternating speed mechanism needs to be fabricated. Gears interconnecting adjacent sleeves can duplicate the motion for the other 3. No reduction on piston skirt size was considered.
For the calculations, it was assumed that the driving mechanism was designed so that the sliding velocity between the piston ring and the sleeve is approximately constant in magnitude and at the value of the peak piston speed of a conventional engine with similar dimensions. This is a conservative figure, since the required speed for fully hydrodynamic lubrication is less, especially at higher engine speeds. Thus, at higher engine speeds, the friction mean effective pressure could be expected high due to the high velocities of the additional moving parts.
Two engines were modeled: one conventional and one designed according to the above recommendations. For both engines the following dimensions were used: Bore=85 mm, Stroke=75 mm, Compression Ratio=10:1, Cylinder Displacement=426 cc, Connecting rod Length=118 mm. A relatively small engine was chosen because this model was derived for small passenger car engines.
In this model, the reciprocating friction was divided into three terms. The first term was the hydrodynamic component mostly from the piston skirt and the connecting rod bearing. The second term was the mixed lubrication term mostly due to ring static tension and therefore is independent of engine load. Finally, the third was the mixed lubrication of the compression rings due to gas loading and was directly proportional to manifold pressure.
In the rotating sleeve case, the two mixed lubrication terms were not included since hydrodynamic lubrication is assumed. The fact that the piston ring lubrication is now hydrodynamic will be accounted by the increase of the hydrodynamic term considered bellow. Furthermore, since the friction in the hydrodynamic lubrication is not as strong a function of the normal force as of sliding speed, the fmep (friction mean effective pressure) on the new design was assumed independent of gas loading and therefore intake manifold pressure.
However, the hydrodynamic term should be higher in the new design in order to account for the extra friction due to sleeve rotation. The hydrodynamic friction term between the rotating sleeve and the piston should be included. Even though this force is indirectly increasing the friction (through the sleeve driving mechanism), for computational purposes, it was assumed directly acting. The overall increased friction was estimated as follows. The hydrodynamic term for the conventional engine was considered as the mean value of an alternating sinusoidal friction of a certain amplitude A. ##EQU1##
By solving the integral, the peak value A was calculated as EQU A=.pi.*fmep/2
Then, the hydrodynamic fmep for the new engine was assumed equal to that peak value since the sliding speed remained always high. The fact that the hydrodynamic term in the engine model included the connecting rod friction as well, made the estimation even more conservative. The fact that at peak piston speed the sleeve still had to slightly rotate should not affect the friction significantly, since the two velocities were normal to each other. Even if the sleeve surface minimum linear speed is 20% of the piston speed at that instant, the resultant vector is less than 2% increased in magnitude. Note that Patton's model includes the ring hydrodynamic portion in these terms. Thus, the overall increase of these terms does take under consideration the increase in the viscous ring friction.
The friction due to the two journal bearings that support the sleeve on the block was also included. A finite element analysis code for journal bearing performance prediction was used. The bearings had to be designed in order to take the side loads transferred from the connecting rod during the power stroke. The two following load criteria derived from the UT Fractal Engine Model had to,be met by the bearings. 50 Atm peak pressure at 20 degrees after top dead center at 1000 rpm and 100 Atm peak pressure (highly exaggerated value to account for a safety factor) at the same crank angle at 3000 rpm. After the design was completed, the sleeve bearings were assumed to rotate at 54% of engine speed for the fmep contribution. This was estimated from the fact that the peak sleeve linear speed should match the peak piston speed. This corresponds to an angular speed equal to the engine speed times stroke to bore ratio. The minimum speed should be 20% of the maximum in order to minimize losses. Finally, the mean would be roughly the average of the two.
The losses from the driving mechanisms were not included. However, gears or chains can be designed to operate at relatively high efficiencies and therefore were not expected to alter the results by much. Graph 4 shows the results obtained. ##STR5##
Graph 5. Piston assembly friction comparison of the proposed design with a conventional engine, high sleeve speed.
Note that for speeds bellow 3600 rpm and medium load, the new design demonstrates lower fmep. The break even point is raised to 4200 rpm for full load. At engine speeds that most vehicles cruise, the fmep reduction is evident.
If the engine is required to operate at higher engine speeds for long periods of time, the sleeves could be geared lower in order to reduce the fmep at these speeds. The trade off could possibly be slightly higher friction and some wear at lower engine speeds as compared to the high sleeve speed scenario, due to the possibility of not having enough sliding speed for a fully hydrodynamic film. The previous assumption that the sleeve speed at TDC needed to be equal to the maximum piston speed in order to retain the fluid lubrication on the compression rings at dead centers is excessive even for medium speeds. Drawings of the sleeve valve test engine used on Ricardo's experiments show that the sleeve linear speed was only a fraction of the peak piston speed and during those tests the engine speed did not exceed 2000 rpm (however, the ring pack used in that engine was of different design compared to modem automotive engines, and the lubricant viscosity was probably high compared to modern energy saving multigrade oils). Furthermore, the numerical solutions of the Reynold's equation discussed earlier indicate that with a conventional ring profile and an SAE20W oil, a sleeve surface speed of 3m/s can create film thickness of well over 0.5 microns at TDC with the typical peak cylinder pressures of spark ignition engnes. With the revised ring profile (FIG. 8), even lower sleeve speeds will suffice. The average sleeve speed is assumed half of the first case, and therefore the increase in the hydrodynamic piston friction is also half. The reason is that since the sleeves are driven by gear mechanisms, the peak sleeve speed is proportional to crankshaft speed. Therefore, the sleeve surface speed at very low rpm and high load may be insufficient for complete asperity contact. Also, the model used to generate the rotating sleeve engine friction ignores mixed lubrication, and thus this potential small increase caused by mixed lubrication does not show in the graph. However, as discussed in a previous section, the low sleeve speed rotating sleeve engine is still expected to show lower friction than conventional engines, even at this operating condition. ##STR6##
Graph 6. Piston assembly friction comparison of the proposed design with a conventional engine, low sleeve speed. Note that in the above graph, the fmep for the new design could be underestimated at low engine speeds.
For engines that operate at fairly narrow rpm range, the sleeve speed could be optimized for wear and friction.
Additional Efficiency Benefits
The rotating sleeve concept can further improve overall engine efficiency indirectly with mechanisms that are not as obvious.
According to Gardner at al. the optimum compression ratio on Direct Injection (DI) diesels is limited by the reduction of mechanical efficiency at high values of compression ratio, mostly due to piston ring friction. In that study, it was shown that the indicated thermal efficiency (engine efficiency disregarding frictional losses) was increasing with increasing compression ratio throughout the range of 13 to 22 to 1 that was tried. However, the friction of the piston rings was also increasing with increasing compression ratio. This effect limited the optimum compression ratio (for best brake thermal efficiency) to a value of 15:1. With the application of the rotating sleeve concept, a severe reduction of piston ring friction will be achieved, and thus, the optimum compression can be higher, achieving even higher thermal efficiency. The higher compression ratio will also help cold starting performance that is usually a problem with diesel engines . According to the same study, satisfactory cold start performance is normally achieved with compression ratios higher than the optimum value for best efficiency.
In conventional engines, the stress on the block caused by the cylinder head bolts as well as thermal stresses, cause liner distortions. According to Lawrence (1990) those distortions can alter the perfectly circular shape of the cross section of the cylinder achieved by the machining process. These distortions cause increased blow-by and oil consumption. Note that this distortion is more pronounced in the upper part of the liner that corresponds to piston locations where the gas pressure and thus the potential for blow by is the highest. In the proposed design, the liner will not be loaded from the cylinder head bolts and thus, free of mechanical distortions. The continuous rotation will eliminate or reduce any temperature gradients around the liner. The perfectly circular liner cross section will result in a more effective gas sealing, and thus reduction of blow-by and oil consumption. Any reduction of blow-by can be considered as direct efficiency benefit.
Conclusion
It has been demonstrated that the new design can provide useful friction reductions and piston ring and liner wear elimination or severe reduction for a variety of engines. More efficient engines of virtually infinite life can be produced. Rotating sleeve engines could be designed in such a way that valve train components (that will still wear with the existing rate) can be rapidly replaced without requiring a complete disassembly. Thus, the engine overhaul cost can be significantly reduced.
Best results are expected on engines that operate at relatively low speed ranges. Then, the motion of the sleeve can be tailored for one speed without excessive hydrodynamic losses (due to excessive sleeve speeds) or boundary friction (due to too low sleeve speeds). However, the durability and frictional reduction requirement could make certain applications more feasible than others. Engines can be designed to operate at high brake mean effective pressures and engine speeds, and thus more power per engine weight without sacrifices in their durability. Modern diesel engine emission requirements demand the use of excessive exhaust gas recirculation that can increase corrosive wear on rings and liners (Needleman and Mandhavan, 1988) increasing the need for anti-corrosion additives. The fluid lubrication will create an immunity of the cylinder wear in the chemical composition of the lubricant. Additives that were specifically formulated for piston ring lubrication may not be necessary on rotating sleeve engines. Thus, the products of the combustion of the lubricant (that happens to some degree in all engines) may not be as toxic as they have to be for conventional engines.